Super-turbocharger having a high speed traction drive and a continuously variable transmission

ABSTRACT

A super-turbocharger utilizing a high speed, fixed ratio traction drive that is coupled to a continuously variable transmission to allow for high speed operation is provided. A high speed traction drive is utilized to provide speed reduction from the high speed turbine shaft. A second traction drive provides infinitely variable speed ratios through a continuously variable transmission. Gas recirculation in a super-turbocharger is also disclosed.

CROSS-REFERENCE TO RELATED APPLICATION

This patent application is a continuation of U.S. application Ser. No.14/055,213, filed Oct. 16, 2013, which application is a divisional ofU.S. application Ser. No. 12/701,440, filed Feb. 5, 2010, whichapplication is a continuation-in-part of U.S. application Ser. No.12/536,421, filed Aug. 5, 2009, which application claims the benefit ofU.S. Provisional Patent Application Ser. No. 61/086,401, filed Aug. 5,2008, the entire teachings and disclosure of which are incorporated byreference thereto.

BACKGROUND

Conventional turbochargers are driven by waste exhaust heat and gases,which are forced through an exhaust turbine housing onto a turbinewheel. The turbine wheel is connected by a common turbo-shaft to acompressor wheel. As the exhaust gases hit the turbine wheel, bothwheels simultaneously rotate. Rotation of the compressor wheel draws airin through a compressor housing, which forces compressed air into theengine cylinder to achieve improved engine performance and fuelefficiency. Turbochargers for variable speed/load applications aretypically sized for maximum efficiency at torque peak speed in order todevelop sufficient boost to reach peak torque. However, at lower speeds,the turbocharger produces inadequate boost for proper engine transientresponse.

To overcome these problems and provide a system that increasesefficiency, a super-turbocharger can be used, which combines thefeatures of a supercharger and a turbocharger. Super-turbochargers mergethe benefits of a supercharger, which is primarily good for high torqueat low speed, and a turbocharger, which is usually only good for highhorsepower at high speeds. A super-turbocharger combines a turbochargerwith a transmission that can put engine torque onto the turbo shaft forsupercharging and elimination of turbo lag. Once the exhaust energybegins to provide more work than it takes to drive the compressor, thesuper-turbocharger recovers the excess energy by applying the additionalpower to the piston engine, usually through the crankshaft. As a result,the super-turbocharger provides both the benefits of low speed with hightorque and the added value of high speed with high horsepower all fromone system.

SUMMARY

An embodiment of the present invention may comprise a super-turbochargerthat is coupled to an engine system comprising: an engine; a turbinethat generates turbine rotational mechanical energy from enthalpy ofexhaust gas produced by the engine; a compressor that compresses intakeair and supplies compressed air to the engine; a shaft having portionsthat are connected to the turbine and the compressor; a traction drivethat is coupled to the shaft that reduces rotational speed of the shaftto a lower rotational speed at an output of the traction drive andtransmits power to and from the shaft; a transmission that connects theoutput of the traction drive and the engine system that transmits thepower between the engine system and the traction drive.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a side view illustration of an embodiment of asuper-turbocharger.

FIG. 2 is a transparency isometric view of the embodiment of thesuper-turbocharger of FIG. 1.

FIG. 3A is a side transparency view of an embodiment of thesuper-turbocharger illustrated in FIGS. 1 and 2.

FIG. 3B is a side cutaway view of another embodiment of asuper-turbocharger.

FIG. 3C is a side transparency view of modifications to the embodimentof the super-turbocharger illustrated in FIGS. 1, 2 and 3A.

FIGS. 4-9 are various drawings of a super-turbocharger using anembodiment of a multi-diameter planetary roller traction drive.

FIG. 10 is an illustration of another embodiment of a high speedtraction drive.

FIGS. 11 and 12 are illustrations of an embodiment of a tractioncontinuously variable transmission.

FIG. 13 is a side cutaway view of another embodiment.

FIG. 14A is a schematic view of an embodiment of a super-turbochargedgas recirculation device.

FIG. 14B is a schematic view of another embodiment of asuper-turbocharged gas recirculation device.

FIG. 14C is a schematic view of another embodiment of asuper-turbocharged gas recirculation device.

FIG. 14D is a graph of valve lift, flow rate and cylinder pressureversus piston position for the embodiments of FIGS. 14A-C.

FIG. 14E is a PV graph of cylinder pressure versus cylinder volume forthe embodiments of FIGS. 14A-C.

FIG. 15 is a graphical illustration of simulated BSFC improvement.

FIG. 16 is a simplified single line form illustration of one embodimentof a high efficiency, super-turbocharged engine system.

FIG. 17 is a detailed diagram of the embodiment of the high efficiencysuper-turbocharged engine system of FIG. 16.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1 is a schematic illustration of an embodiment of asuper-turbocharger 100 that uses a high speed traction drive 114 and acontinuously variable transmission 116. As shown in FIG. 1, thesuper-turbocharger 100 is coupled to the engine 101. Thesuper-turbocharger includes a turbine 102 which is coupled to engine 101by an exhaust conduit 104. The turbine 102 receives the hot exhaustgases from the exhaust conduit 104 and generates rotational mechanicalenergy prior to exhausting the exhaust gases in an exhaust outlet 112. Acatalyzed diesel particulate filter (not shown) can be connected betweenthe exhaust conduit 104 and turbine 102. Alternatively, the catalyzeddiesel particulate filter (not shown) can be connected to the exhaustoutlet 112. The rotational mechanical energy generated by the turbine102 is transferred to the compressor 106 via a turbine/compressor shaft,such as shaft 414 of FIG. 4, to rotate a compressor fan disposed in thecompressor 106, which compresses the air intake 110 and transmits thecompressed air to a conduit 108, which is coupled to an intake manifold(not shown) of the engine 101. As disclosed in the above referencedapplication, super-turbochargers, unlike turbochargers, are coupled to apropulsion train to transfer energy to and from the propulsion train.The propulsion train, as referred to herein, may comprise the engine101, the transmission of a vehicle in which the engine 101 is disposed,the drive train of a vehicle in which the engine 101 is disposed, orother applications of the rotational mechanical energy generated byengine 101. In other words, rotational mechanical energy can be coupledor transferred from the super-turbocharger to the engine through atleast one intermediate mechanical device such as a transmission or drivetrain of a vehicle, and vice versa. In the embodiment of FIG. 1, therotational mechanical energy of the super-turbocharger is coupleddirectly to a crankshaft 122 of engine 101 through a shaft 118, a pulley120 and a belt 124. As also illustrated in FIG. 1, a high speed tractiondrive 114 is mechanically coupled to a continuously variabletransmission 116.

In operation, the high speed traction drive 114, of FIG. 1, is a fixedratio, high speed traction drive that is mechanically coupled to theturbine/compressor shaft using a traction interface to transferrotational mechanical energy to and from the turbine/compressor shaft.The high speed traction drive 114 has a fixed ratio which may differ inaccordance with the size of the engine 101. For small engines, a largefixed ratio of the high speed traction drive 114 is required.

For smaller engines, the compressor and turbine of a super-turbochargermust necessarily be smaller to maintain a small engine size and to matchthe flow requirements of the compressor and turbine. In order for asmaller turbine and a smaller compressor to function properly, they haveto spin at a higher rpm. For example, smaller engines may require thecompressor and turbine to spin at 300,000 rpm. For very small engines,such as half liter engines, the super-turbocharger may need to spin at900,000 rpm. One of the reasons that smaller engines require compressorsthat operate at a higher rpm level is to avoid surge. In addition, tooperate in an efficient manner, the tip velocity of the compressor mustbe just short of the speed of sound. Since the tips are not as long insmaller compressors, the tips of a smaller compressor are not moving asfast as the tips on larger compressors at the same rpm. As the size ofthe compressor decreases, the rotational speed required to operateefficiently goes up exponentially. Since gears are limited toapproximately 100,000 rpm, standard gear systems cannot be used toachieve the power take off at the higher speeds necessary for a carengine super-turbocharger. Therefore, various embodiments use a highspeed traction drive 114 to add and receive power from the turbo shaft.

The rotational mechanical energy from the high speed traction drive 114is therefore reduced to an rpm level that is variable depending upon therotational speed of the turbine/compressor, but at an rpm level that iswithin the operating range of the continuously variable transmission(CVT) 116. For example, the high speed traction drive 114 may have anoutput that varies between zero and 7,000 rpm while the input from theturbine/compressor shaft may vary from zero to 300,000 rpm, or greater.The continuously variable transmission 116 adjusts the rpm level of thehigh speed traction drive 114 to the rpm level of the crankshaft 122 andpulley 120 to apply rotational mechanical energy to engine 101, orextract rotational mechanical energy from engine 101 at the proper rpmlevel. In other words, the continuously variable transmission 116comprises an interface for transferring rotational mechanical energybetween engine 101 and the high speed traction drive 114 at the properrpm level which varies in accordance with the engine rotational speedand the turbine/compressor rotational speed. Continuously variabletransmission 116 can comprise any desired type of continuously variabletransmission that can operate at the required rotational speeds and havea ratio to match the rotational speed of the crankshaft 122 or othermechanisms coupled, directly or indirectly, to the engine 101. Forexample, in addition to the embodiments disclosed herein, two rollerCVTs can be used as well as traction ball drives and pushing steel beltCVTs.

An example of a continuously variable transmission that is suitable foruse as continuously variable transmission 116, disclosed in FIG. 1, isthe continuously variable transmission disclosed in FIGS. 11 and 12.Other examples of continuously variable transmissions that can be usedas the continuously variable transmission 116 of FIG. 1 include U.S.Pat. No. 7,540,881 issued Jun. 2, 2009, to Miller et al. The Millerpatent is an example of a traction drive, continuously variabletransmission that uses a planetary ball bearing. The traction drive ofMiller is limited to about 10,000 rpm so that the Miller continuouslyvariable transmission is not usable as a high speed traction drive, suchas high speed traction drive 114. However, the Miller patent doesdisclose a continuously variable transmission that uses a traction driveand is suitable for use as an example of a continuously variabletransmission that could be used as continuously variable transmission116 as illustrated in FIGS. 1-3. Another example of a suitablecontinuously variable transmission is disclosed in U.S. Pat. No.7,055,507 issued Jun. 6, 2006, to William R. Kelley, Jr., and assignedto Borg Warner. Another example of a continuously variable transmissionis disclosed in U.S. Pat. No. 5,033,269 issued Jul. 23, 1991 to Smith.Further, U.S. Pat. No. 7,491,149 also discloses a continuously variabletransmission that would be suitable for use as continuously variabletransmission 116. U.S. Pat. No. 7,491,149 issued Feb. 17, 2009 toGreenwood et al. and assigned to Torotrak Limited discloses an exampleof a continuously variable transmission that uses a traction drive thatcan be used as the continuously variable transmission 116. All of thesepatents are specifically incorporated by reference for all that theydisclose and teach. European Application No. 92830258.7, published Aug.9, 1995, as Publication No. 0517675B1, also illustrates anothercontinuously variable transmission 3 that is suitable for use as thecontinuously variable traction drive 116.

Various types of high speed traction drives can be used as the highspeed traction drive 114. For example, the high speed planetary tractiondrive 406 disclosed in FIGS. 4-9 and the high speed planetary drive ofFIG. 10 can be used as high speed traction drive 114.

Examples of high speed drives that use gears are disclosed in U.S. Pat.No. 2,397,941 issued Apr. 9, 1946 to Birgkigt and U.S. Pat. No.5,729,978 issued Mar. 24, 1998 to Hiereth et al. Both of these patentsare specifically incorporated herein by reference for all that theydisclose and teach. Both of these references use standard gears and donot use traction drives. Hence, even with highly polished, speciallydesigned gearing systems, the gears in these systems are limited torotational speeds of approximately 100,000 rpm or less. U.S. Pat. No.6,960,147 issued Nov. 1, 2005 to Kolstrup and assigned to RuloundsRoadtracks Rotrex A/S discloses a planetary gear that is capable ofproducing gear ratios of 13:1. The planetary gear of Kolstrup is anexample of a high speed drive that could be used in place of a highspeed traction drive 114 of FIG. 1. U.S. Pat. No. 6,960,147 is alsospecifically incorporated herein by reference for all that it disclosedand teaches.

FIG. 2 is a schematic side transparency view of the super-turbocharger100. As shown in FIG. 2, turbine 102 has an exhaust conduit 104 thatreceives exhaust gases that are applied to the turbine fan 130.Compressor 106 has a compressed air conduit 108 that supplies compressedair to the intake manifold. Compressor housing 128 encloses thecompressor fan 126 and is coupled to the compressed air conduit 108. Asdisclosed above, high speed traction drive 114 is a fixed ratio tractiondrive that is coupled to a continuously variable transmission 116. Thecontinuously variable transmission 116 drives shaft 118 and pulley 120.

FIG. 3A is a side transparency view of the embodiment of thesuper-turbocharger 100 illustrated in FIGS. 1 and 2. Again, as shown inFIG. 3A, turbine 102 includes a turbine fan 130, while compressor 106includes a compressor fan 126. A shaft (not shown) connecting theturbine fan 130 and compressor fan 126 is coupled to a high speedtraction drive 114. Rotational mechanical energy is transferred from thehigh speed traction drive 114 to a transfer gear 132 that transfers therotational mechanical energy to a CVT gear 134 and the continuouslyvariable transmission (CVT) 116. The continuously variable transmission116 is coupled to the shaft 118 and pulley 120.

FIG. 3B is a schematic cutaway view of another example of asuper-turbocharger 300 that is coupled to an engine 304. As shown inFIG. 3B, the turbine 302 and the compressor 306 are mechanically coupledby shaft 320. High speed traction drive 308 transfers rotationalmechanical energy to, and receives rotational mechanical energy from,transfer gear 322. A specific example of a high speed traction drive 308is illustrated in FIG. 3B. Transfer gear 322 transfers rotationalmechanical energy between the traction drive 308 and the continuouslyvariable transmission 310. A specific example of a continuously variabletransmission 310 is also illustrated in FIG. 3B. Shaft 312, pulley 314and belt 316 transfer rotational mechanical energy between thecrankshaft 318 and the continuously variable transmission 310.

FIG. 3C is a side schematic cutaway view of modifications to theembodiment of the super-turbocharger 100 illustrated in FIGS. 1, 2 and3A. As shown in FIG. 3C, turbine 102 and compressor 106 are coupledtogether by a shaft (not shown). High speed traction device 114 iscoupled to the shaft. Rotational mechanical energy is transferred fromthe high speed traction device 114 to a transfer gear 132 that transfersthe rotational mechanical energy to transmission gear 134. High speedtraction drive 114, transfer gear 132 and transmission gear 134 may allbe housed in the same housing. Transmission gear 134 is connected to atransmission 140 that can comprise a manual gear box, a CVT, a straightshaft, an automatic gear box, or a hydraulic transmission. Transmission140 is then connected to a shaft 118 which is connected to a pulley 120.Pulley 120 is coupled to the propulsion train. In an alternativeembodiment, pulley 120 is coupled to an electric notor/generator 142.

FIG. 4 is a schematic transparency view of another embodiment ofsuper-turbocharger 400 that utilizes a high speed traction drive 416that is coupled to a continuously variable transmission 408. As shown inFIG. 4, the turbine 404 is mechanically coupled to the compressor 402with a compressor/turbine shaft 414. Rotational mechanical energy istransferred between the compressor/turbine shaft 414 and themulti-diameter traction drive 416 in the manner disclosed in more detailbelow. Transfer gear 418 transfers rotational mechanical energy betweenthe multi-diameter traction drive 416 and the CVT gear 420 of thecontinuously variable transmission 408. Shaft 410 and pulley 412 arecoupled to the continuously variable transmission 408 and transfer powerbetween the continuously variable transmission 408 and a propulsiontrain.

FIG. 5 is a side cutaway schematic view of the multi-diameter tractiondrive 416 that is coupled to the transfer gear 418, which is in turncoupled to the CVT gear 420. The compressor/turbine shaft 414 has apolished, hardened surface on a central portion, as disclosed in moredetail below, that functions as a sun drive in the multi-diametertraction drive 416.

FIG. 6 is an exploded view 600 of the embodiment of thesuper-turbocharger 400 illustrated in FIG. 4. As shown in FIG. 6,turbine housing 602 houses a turbine fan 604. The hot side cover plate606 is mounted adjacent to the turbine fan 604 and the main housingsupport 608. A ring seal 610 seals the exhaust at the hot side coverplate 606. Ring roller bearing 612 is mounted in the ring roller 614.Compressor/turbine shaft 414 extends through the main housing support608. The hot side cover plate 606 connects with the turbine fan 604.Planet carrier ball bearing 618 is mounted on the planet carrier 620.Multi-diameter ring rollers 622 are rotationally connected to the planetcarrier 620. Oil feed tubes 624 are used to supply traction fluid to thetraction surface. Planet carrier 626 is mounted to the planet carrier620 and uses a planet carrier ball bearing 628. Fixed ring 630 is thenmounted outside of planet carrier 626. Cage 632 is mounted between thefixed ring 630 and the cool side cover plate 636. Compressor fan 638 iscoupled to the compressor/turbine shaft 414. Compressor housing 640encloses the compressor fan 638. The main housing support 608 alsosupports the continuously variable transmission and the transfer gear418. Various bearings 646 are used to mount the transfer gear 418 andthe main housing support 608. The continuously variable transmissionincludes a CVT cover 642 and a CVT bearing plate 644. CVT gear 420 ismounted inside the main housing support 608 with bearings 650. CVTbearing plate 652 is mounted on the opposite side of the CVT gear 420from the CVT bearing plate 644. CVT cover 654 covers the variousportions of the CVT device. Shaft 410 is coupled to the continuouslyvariable transmission. Pulley 412 is mounted on shaft 410 and transfersrotational mechanical energy between shaft 410 and a propulsion train.

FIG. 7 is a perspective view of isolated key components of themulti-diameter traction drive 416, as well as the turbine fan 604 andcompressor fan 638. As shown in FIG. 7, the compressor/turbine shaft 414is connected to the turbine fan 604 and compressor fan 638, and passesthrough the center of the multi-diameter traction drive 416. Themulti-diameter traction drive 416 includes multi-diameter planet rollers664, 666 (FIG. 9), 668. These multi-diameter planet rollers arerotationally coupled to a planet carrier 626 (FIG. 9). Balls 656, 658,660, 662 rest on an incline surface for ball ramps on the fixed ring630. Ring roller 614 is driven by an inner diameter of themulti-diameter planet rollers 664, 666, 668, as disclosed in more detailbelow.

FIG. 8 is a side cutaway view of the multi-diameter traction drive 416.As shown in FIG. 8, the compressor/turbine shaft 414 is hardened andpolished to form a traction surface that is used as a sun roller 674that has a traction interface 676 with the multi-diameter planet roller664. The multi-diameter planet roller 664 rotates along themulti-diameter planet roller axis 672. The multi-diameter planet roller664 contacts the fixed ring 630 at the interface 690 of the planetroller 664 and the fixed ring 630. The multi-diameter planet roller 664contacts the ring roller 614 at interface 691, which is a differentradial distance from the multi-diameter planet roller axis 672, than theinterface 691. FIG. 8 also illustrates the planet carrier 626 and theball ramp 630 that intersects with the ball 656, and ball ramp 631 thatintersects with ball 660. The balls 656, 658, 660, 662 are wedged inbetween a housing (not shown) and the ball ramp, such as ball ramp 630,on the fixed ring 664. When torque is applied to the ring roller 614,this causes the fixed ring 664 to move slightly in the direction of therotation of the ring roller 614. This causes the balls to move up thevarious ball ramps, such as ball ramps 630, 631, which in turn causesthe fixed ring 630 to press against the multi-diameter planet rollers664, 666, 668. Since the interface 691 of the planet roller 664 andfixed ring 630 is sloped, and the interface of the planet roller 664 andring roller 690 is sloped, an inward force on the multi-diameter planetroller 664 is generated, which generates a force on the tractioninterface 676 to increase the traction at the traction interface 676between the multi-diameter planet roller 664 and the sun roller 674. Inaddition, a force is created at the interface 691 of the multi-diameterplanet roller 664 and the ring roller 614, which increases traction atinterface 691. As also shown in FIG. 8, the compressor fan 630 and theturbine fan 604 are both coupled to the compressor/turbine shaft 414.Ring roller 614 is coupled to the transfer gear 418, as also shown inFIG. 8.

FIG. 9 is a side cutaway view of the multi-diameter traction drive 416.As shown in FIG. 9, the sun roller 674 rotates in a clockwise direction,as shown by rotation direction 686. The multi-diameter planet rollers664, 666, 668 have outer diameter roller surfaces, such as outerdiameter roller surface 688 of multi-diameter planet roller 664. Theseouter diameter roller surfaces contact the sun roller 674 which causethe multi-diameter planet rollers 664, 666, 668 to rotate in acounter-clockwise direction, such as rotational direction 684 ofmulti-diameter planet roller 666. The multi-diameter planet rollers 664,666, 668 also have an inner diameter roller surface, such as innerroller diameter roller surface 680 of multi-diameter planet roller 664.The inner diameter roller surface of each multi-diameter planet rollercontacts the roller surface 687 of the ring roller 614. Hence, theinterface 678 of planet roller 664 with the roller surface 687 of ringroller 614 constitutes a traction interface that transfers rotationalmechanical energy when a traction fluid is applied. The interfacebetween each of the multi-diameter planet rollers 664, 666, 668 and sunroller 674 also constitutes a traction interface that transfersrotational mechanical energy upon application of a traction fluid.

As indicated above with respect to FIGS. 8 and 9, the fixed ring 630generates a force, which pushes the multi-diameter planet rollers 664,666, 668 towards the sun roller 674 to generate traction. Each of themulti-diameter planet rollers 664, 666, 668 is rotationally attached tothe planet carrier 626 with planet roller axes, such as themulti-diameter planet roller axis 672 of the multi-diameter planetroller 664. These axes have a slight amount of play so that themulti-diameter planet rollers 664, 666, 668 can move slightly and createa force between the sun roller 674 and the outer diameter roller surfaceof the multi-diameter planet roller 664, 666, 668, such as the outerdiameter of the roller surface 688 of the planet roller 664. Themovement of the multi-diameter planet roller 664 towards the sun roller674 also increases the traction at the interface of the multi-diameterplanet rollers 664, 666, 668 and the ring roller 614, since theinterface between the multi-diameter planet rollers 664, 666, 668 andthe ring roller 614, such as interface 678, is sloped. The contact withthe multi-diameter planet rollers 664, 666, 668 with the roller surface687 of ring roller 614 causes the planet carrier 626 to rotate in aclockwise direction, such as the rotational direction 682, illustratedin FIG. 9. As a result, the ring roller 614 rotates in acounter-clockwise direction, such as rotational direction 687, anddrives the transfer gear 418 in a clockwise direction.

FIG. 10 is a schematic cross sectional view of another embodiment of ahigh speed traction drive 1000. As shown in FIG. 10, a shaft 1002, whichis a shaft, that connects a turbine and a compressor insuper-turbocharger, can act as a sun roller in the high speed tractiondrive 1000. Planet roller 1004 contacts the shaft 1002 at tractioninterface 1036. Planet roller 1004 rotates on an axis 1006 usingbearings 1008, 1010, 1012, 1014. As also shown in FIG. 10, gear 1016 isdisposed and connected to the outer surface of the carrier 1018. Carrier1018 is coupled to a housing (not shown) via bearings 1032, 1034, whichallow the carrier 1018 and gear 1016 to rotate. Fixed rings 1020, 1022include ball ramps 1028, 1030, respectively. Ball ramps 1028, 1030 aresimilar to the ball ramps 630 illustrated in FIGS. 7 and 8. As the gear1016 moves, the balls 1024, 1026 move in the ball ramps 1028, 1030,respectively, and force the fixed rings 1020, 1022 inwardly towards eachother. A force is created between the fixed rings 1020, 1022 and thesurface of the planet roller 1004 at traction surfaces 1038, 1040 as theballs 1024, 1026 force the fixed ramps 1020, 1022 inwardly towards eachother. The force created by the fixed rings 1090, 1092 also forces theplanet roller 1004 downwardly, as illustrated in FIG. 10, so that aforce is created between the shaft 1002 and the planet roller 1004 atthe traction interface 1036. As a result, greater traction is achievedat a traction interface 1036 and the traction surfaces 1038, 1040.Traction fluid is applied to these surfaces, which becomes sticky andincreases friction at the traction interfaces, as the traction fluid isheated as a result of the friction created at the traction interfaces1036, 1038, 1040.

The high speed traction drive 1000, illustrated in FIG. 10, is capableof rotating at high speeds in excess of 100,000 rpm, which isunachievable by gearing systems. For example, the high speed tractiondrive 1000 may be able to rotate at speeds greater than 300,000 rpm.However, high speed traction drive 1000 is limited to a gear ratio ofapproximately 10:1 because of the physical limitations of size. The highspeed traction drive 1000 may utilize three planet rollers, such asplanet roller 1006 that are disposed radially around the shaft 1002. Asillustrated in FIG. 9, the size of the planet rollers is limited withrespect to the sun roller. If the diameter of the planet rollers in FIG.9 increases, the planet rollers will abut each other. Hence, gear ratiosof only about 10:1 can be reached with a planetary traction drive, suchas illustrated in FIG. 10, while the multi-diameter planet drives thatare connected to a planet carrier, such as illustrated in FIGS. 7-9, mayhave ratios of as much as 47:1 or greater. Accordingly, if a compressoris required for a smaller engine that must rotate at 300,000 rpm to beefficient, a 47:1 ratio traction drive, such as illustrated in FIGS.7-9, can reduce the maximum rotational speed of 300,000 rpm toapproximately 6,400 rpm. Standard geared or traction continuouslyvariable transmissions can then be used to transfer the rotationalmechanical energy between the high speed traction drive and thepropulsion train of the engine.

As disclosed above, the high speed traction drive 1000, illustrated inFIG. 10, may have a ratio as large as 10:1. Assuming a rotational speedof the shaft 1002 is 300,000 rpm for a super-turbocharger for a smallengine, the 300,000 rpm rotational speed of the shaft can be reduced to30,000 rpm at gear 1016. Various types of continuously variabletransmissions 116 can be used that operate up to 30,000 rpm usingstandard gearing techniques. Traction drive continuously variabletransmissions, such as the traction drive continuously variabletransmission illustrated in FIGS. 11 and 12, can also be used as thecontinuously variable transmission 116, illustrated in FIG. 1. Further,ratios of up to 100:1 may be achievable with the multi-diameter tractiondrive 416, illustrated in FIG. 4-9. Accordingly, small engines of 0.5liters, which may require a compressor that operates at 900,000 rpm, canbe reduced to 9,000 rpm, which is a rotational speed that can be easilyutilized by various continuously variable transmissions 116 to couplerotational mechanical energy between a propulsion train and aturbine/compressor shaft.

FIGS. 11 and 12 illustrate an example of a continuously variabletraction drive transmission that can be used as the continuouslyvariable transmission 116 of FIG. 1. The traction drive continuouslyvariable transmission illustrated in FIGS. 11 and 12 operates bytranslating races 1116, 1118 in a lateral direction on race surfacesthat have a radius of curvature that causes contact locations of theball bearings to move, which, in turn, causes the balls to rotate with adifferent spin angle to drive race 1122 at different speeds. In otherwords, the contact location of each of the bearings on the race surfacesis changed as a result of the lateral translation of the races 1116,1118, which alters the speed at which the bearings are rotating at thecontact location, as explained in more detail below.

As shown in FIG. 11, input shaft 1102 is coupled to the transfer gear132 (FIG. 3A). For example, splines 1104 may be splined to the CVT gear134, illustrated in FIG. 3A. Hence, the spline input gear 1104 of theinput shaft 1102 can be coupled to the super-turbocharger through a highspeed traction drive 114, as illustrated in FIG. 3A. In this manner,input torque from the propulsion train is used to drive the spline inputgear 1104 of the input shaft 1102. The input torque on the spline inputgear 1104 imparts a spin in rotational direction 1112 on both the inputshaft 1102 and its associated structure including input race 1114. Inputrace 1116 is also spun around the axis of rotation 1106 in response tothe torque imparted by spline 1166 from the input shaft 1102 to theinput race 1116. The rotation of the input shaft 1102, input race 1114and input race 1116 impart a spin on the plurality of ball bearings 1132because the stationary race 1120 impedes the rotation of the ballbearings at the contact point with stationary race 1120. Input race 1114and input race 1116 rotate at the same angular speed since they arecoupled together through spline 1116. Input race 1114 and input race1116 cause the ball bearings 1132 to spin in a substantially verticalorientation since the ball bearings 1132 contact the stationary race1120. The contact of the ball bearings 1132 against the stationary race1120 also causes the ball bearings 1132 to precess around the perimeterof the races 1114, 1116, 1118, 1120. In the embodiment illustrated inFIG. 11, there may be as many as 20 ball bearings 1132 that rotate onthe surfaces of the races 1114, 1116, 1118, 1120. The rotation of theball bearings 1132 as a result of being driven by input race 1114 andinput race 1116 creates a tangential contact of the ball bearings 1132on the output race 1118. Depending upon the contact position of the ballbearings 1132 on the output race 1118, the ratio of the rotational speedof the input races 1114, 1116 with respect to the output race 1118 canbe varied. Output race 1118 is coupled to output gear 1122. Output gear1122 engages output gear 1124, which in turn is connected to the outputshaft 1126.

The manner in which the traction drive continuously variabletransmission 1100, illustrated in FIG. 11, shifts the ratio between theinput shaft 1102 and the output shaft 1126 is accomplished by changingthe relative position of the contact point between the four races 1114,1116, 1118, 1120 that are in contact with the ball bearings 1132. Themanner in which the contact surfaces of the races 1114, 1116, 1118, 1120with the ball bearings 1132 is changed is by shifting the position ofthe translating clamp 1152. The translating clamp 1152 is movedhorizontally, as illustrated in FIG. 11, in response to electricactuator 1162. Electric actuator 1162 has a shaft that engages thetelescopic shifter 1158 and rotates the telescopic shifter 1158.Telescopic shifter 1158 has different thread types on an inside portionand an outside portion. A difference in thread pitch of the differentthread types causes the translating clamp 1152 to translate horizontallyin response to rotation of the shaft of the electric actuator 1162,which imparts rotation in the telescopic shifter 1158. Lateraltranslation of the translating clamp 1152, which is in contact withbearing clamp 1164, causes lateral transition of input race 1116 andoutput race 1118. Lateral translation of the input race 1116 and outputrace 1118 may vary, in the embodiment illustrated in FIG. 11, byapproximately one-tenth of an inch. The translation of the input race1116 and the output race 1118 changes the angle of contact between theball bearings 1132 and the output race 1118, which changes the ratio, orspeed at which the ball bearings 1132 are moving in the races because ofa change in contact angle between the stationary race 1120 and inputrace 1114 and input race 1116. The combination of the change in anglebetween the races allows the contact velocity, or the point of contactbetween the ball bearings 1132 and output race 1118, to vary whichresults in a variation of speed of between 0 percent of the rotationalspeed of the input shaft 1102 up to 30 percent of the rotational speedof the input shaft 1102. The variation of speed in the output race 1118of 0 percent to 30 percent of the rotational speed of the input shaft1102 provides a wide range of adjustable rotational speeds that can beachieved at the output shaft 1126.

To ensure proper clamping of the ball bearings 1132 between the races1114, 1116, 1118, 1120, springs 1154, 1156 are provided. Spring 1154generates a clamping force between input race 1114 and stationary race1120. Spring 1156 generates a clamping force between input race 1116 andoutput race 1118. These clamping forces against the ball bearings 1132are maintained over the entire translating distance of the translatingclamp 1152. The telescopic shifter 1158 has threads on an inside surfacethat connect to the threads on the fixed threaded device 1160. The fixedthreaded device 1160 is fixed to housing 1172 and provides a fixedposition relative to the housing 1172 so that the translating clamp 1152is able to translate in a horizontal direction as a result of thedifferential threads on the two sides of the telescopic shifter 1158.

As also illustrated in FIG. 11, the rotating components of the tractiondrive continuously variable transmission 1100 all rotate in the samedirection, i.e. rotational direction 1112 and output rotation 1128 ofthe output gear 1122. Clamping nut 1168 holds spring 1156 in place andpreloads the spring 1156 to create the proper diagonal pressure betweenstationary race 1120 and input race 1114. When the translating clamp1152 is horizontally translated, as illustrated in FIG. 11, there is aslight translation of the input shaft 1102 based upon the angles of theraces 1114-1120 that contact the ball bearings 1132. The spline inputgear 1104 allows translational movement in directions 1108, 1110 basedupon the points at which the ball bearings 1132 contact the races1114-1120 and the particular contact angle of the races with respect tothe ball bearings 1132. Housing 1170 is bolted tightly to housing 1172to contain the spring 1154, which creates the proper amount of clampingforce between input race 1114 and stationary race 1120. Ball bearings1132, as illustrated in FIG. 11, have a rotational progression 1131 inthe four races 1114, 1116, 1118, 1120. The rotational direction 1112 ofthe shaft 1102 causes the gear 1122 to rotate in a rotational direction1128, as illustrated in FIG. 11.

FIG. 12 is a closeup view of the races 1114-1120 and ball 1132,illustrating the operation of the traction drive continuously variabletransmission 1100. As shown in FIG. 12, race 1114 forcibly contacts ball1132 at contact location 1134. Race 1116 forcibly contacts ball 1132 atcontact location 1136. Race 1118 forcibly contacts ball 1132 at contactlocation 1138. Race 1120 forcibly contacts ball 1132 at contact location1140. Each of the contact locations 1134, 1136, 1138, 1140 is located ona common great circle on the surface of the ball 1132. The great circleis located in a plane that contains the center of the ball 1132 and theaxis 1106 of the shaft 1102. Ball 1132 spins about a spin axis 1142passing through the center of the ball 1132 and bisects the great circlecontaining contact locations 1134, 1136, 1138, 1140. The spin axis 1142of the ball 1132 is inclined at an angle 1146 with the vertical axis1144. The inclination angle 1146 is the same for each of the ballsdisposed in the races around the circumference of the traction drive1100. The inclination angle 1146 establishes a mathematical relationshipbetween a distance ratio and a circumferential velocity ratio. Thedistance ratio is the ratio between the first distance 1148, which isthe orthogonal distance from the spin axis 1142 to the contact location1134, and a second distance 1150, which is the orthogonal distance fromthe spin axis 1142 to contact location 1136. This distance ratio isequal to the circumferential velocity ratio. The circumferentialvelocity ratio is the ratio between the first circumferential velocityand the second circumferential velocity, where the first circumferentialvelocity is the difference between the circumferential velocity of ball1132 at race 1114 and a common orbital circumferential velocity of ball1132 and the other balls in the races, while the second circumferentialvelocity is the difference between the circumferential velocity of theball 1132 on the race 1116 and the common orbital circumferentialvelocity of the ball 1132, as well as the other balls disposed in theraces. The radius of curvature of each of the races 1114-1120 is largerthan the radius of curvature of ball 1132. In addition, the radius ofcurvature of each of the races 1114-1120 need not be a constant radiusof curvature, but can vary. Further, the radius of curvature of each ofthe four races does not have to be equal.

When races 1116, 1118 translate simultaneously in a lateral direction,such as lateral translation direction 1108, the speed ratio of therotation of shaft 1102 and the rotational direction 1112 change withrespect to the rotation of the gear 1122 and rotational direction 1128.Translation of races 1116, 1118 in lateral translation direction 1108causes the first distance 1148 to be larger and the second distance 1150to be smaller. Hence, the ratio of distances, as well as thecircumferential velocity ratio, changes, which changes the rotationalspeed of the gear 1122 with respect to shaft 1102.

As indicated above, the continuously variable transmission output is ingear contact with the traction drive speed reduction mechanism thatconnects to the turbine compressor shaft. As indicated above, there areat least two or three different types of traction drive speed reductionsystems that may be used. The typical type is a planetary type tractiondrive for high speed reduction, which is disclosed in FIGS. 6-9, andFIG. 10. If a large speed differential between the turbine shaft and theplanetary roller is desired, the embodiment of FIG. 10 may utilize onlytwo rollers instead of three, in order to get the gear ratio change thatis desired.

With three rollers, a limit of about a 10:1 reduction in speed existsand there may be a need for more like a 20:1 transmission to get thehigh speed 250,000 rpm operation below the 25,000 rpm to which a 10:1transmission would require. Therefore, a two roller planetary tractiondrive can be used in place of a three planetary drive system, in FIG.10, in order to achieve the speed reduction required of the smallesthighest speed systems. Two rollers also provide for lower inertia, aseach roller adds some amount of inertia to the system. For the lowestinertia, two rollers should be sufficient. The width of the tractionroller is slightly wider than a three roller embodiment.

The multi-diameter planet rollers that roll against the shaft are madeof a springy material, e.g., either a spring steel or another material,that allows some deformation of the roller within the outer drum. Theapplication of a spring loaded roller can provide the necessary pressureon the shaft, but not restrict the shaft's ability to find its idealcenter of rotation.

When a turbocharger operates at extremely high speeds, it has balanceconstraints that cause the shaft to need to find its own center ofrotation. The balance will be compensated by the movement of the centershaft. This movement can be compensated by spring-loaded rollers. Thespring-loaded rollers can also be made extremely light weight by makingthem out of a thin band of steel that allows them to operate against theshaft with very low inertia. The band thickness must be thick enough toput sufficient pressure on the traction surfaces to provide the normalforce needed for traction. A cam follower can be disposed inside theroller that will position each roller and hold that position within thesystem. Rollers need to operate in a very straight alignment between theouter drum and the turbine/compressor shaft, but the key to low inertiais lightweight. One or two cam followers can be utilized to hold thesteel band in place, such that the steel band stays in alignment in thesystem.

The ring roller 614 is connected to a gear on the outside surface sothat the ring roller can transmit the power in or out of themulti-diameter traction drive 416. The ring roller 614 can be made innumerous ways. Ring roller 614 can simply be a solid piece of steel orother appropriate material that is capable of transmitting the torque inand out of the multi-diameter traction drive 416. Ring roller 614 can bemade of numerous materials that allow ring roller 614 to be lightweight,but ring roller 614 has to be from a material that can be used as atraction drive surface on the roller surface 687. A proper rollersurface 687 allows the planet rollers 664, 666, 668 to transmit thetorque through traction.

Also, turbine/compressor shaft 414 needs to be held in very accuratealignment. The alignment of the turbine/compressor shaft 414, within thehousing, allows the clearances to be held between the tips of the bladesof the compressor and the compressor housing. A tighter clearanceincreases the compressor efficiency. A more accurate position decreasesthe chance of touching between the turbine compressor fan 638 and thecompressor housing 640. A method of controlling the thrust load thatcomes from compressing the gas against the compressor wheel is necessaryto ensure that there is a minimum of clearance. This can be done using athrust bearing (not shown) that is oil fed or a thrust bearing that is aball bearing or roller bearing type of bearing.

Typically, in a turbocharger, the bearings are, for reliabilitypurposes, sleeve bearings that have an oil clearance both on the insideand the outside in order to allow for the turbine shaft to center itselfin its harmonic rotation. The balancing requirements for a high volumemanufactured turbocharger are reduced by using a double clearancebearing. These bearing types have been used because of the requirementof tighter clearances and more accurate alignment of the shaft of theturbocharger. A ball bearing is used for both holding the compressor andturbine and for maintaining better alignment to the housing from aside-to-side motion perspective. This can be accomplished with one ortwo ball bearings. Alignment of bearings within an outer area that ispressurized with oil allows the bearings to float and allows the bearingto find a center. This does affect the clearance between the housing,turbine and compressor outside edges, but allows thrust clearance toremain small. Turbo shaft bearings provide a third point of constraintto maintain alignment of the rollers. Cam followers in the middle of therollers can keep the rollers at 120 degrees from one another. Two smallcam followers can be used for each roller to eliminate backlash whenpower changes direction.

Also, a larger turbine can be used. The turbine wheel can be made largerin diameter than normal. It is possible to make the turbine outerdiameter even larger than the compressor wheel, without hitting thecritical speed where tips come close to the speed of sound, because thedensity of the exhaust is lower than inlet air and therefore the speedof sound is higher. This allows the exhaust to generate more torque onthe turbine/compressor shaft without higher backpressure. Having highertorque causes the turbine to recover more energy than is required tocompress the intake air. This produces more energy than can be recoveredand transmitted to the engine. More energy from the same exhaust gasflow that is not needed for compression gets transferred to thecrankshaft and creates lower fuel consumption.

Further, turbine efficiency can be improved by using guide vanes thatcontrol the angle of incidence which exhaust gases impact the turbinewheel. This makes the peak efficiency higher, but narrows the speedrange upon which that efficiency is achieved. A narrow speed range isbad for a normal turbocharger, and is not a problem for asuper-turbocharger where the governor can provide the necessary speedcontrol.

Higher backpressure across the turbine compared to the pressure acrossthe compressor can also create an unbalanced super-turbocharger. For anormal turbocharger, this pressure difference is the other way around.Having higher backpressure causes the turbine to recover more energythan is required to compress the intake air. This produces more energythat can be recovered and transmitted to the engine. Higher backpressureis needed for high pressure EGR loops on diesel engines. Highbackpressure normally requires a valve or a restriction, so highbackpressure is normally lost energy because a normal turbochargercannot be unbalanced without over-speeding. Increasing backpressure isbad for gasoline and natural gas engines, because it increases theamount of exhaust gas that gets trapped in the cylinder, which makes theengine more likely to have detonation problems.

In accordance with another embodiment, a second turbine wheel can bepositioned on the turbine/compressor shaft to increase the energyrecovered by the turbine and improve the fuel efficiency of the enginesystem. Also, a second compressor wheel can be positioned on the sameshaft to increase the boost pressure potential of the super-turbochargerand allow intercooling between the stages. This makes the intaketemperature cooler for a given boost and therefore lowers NOx.

In addition, turbine blade cooling can be provided through the wing tipsto reduce temperatures in high temperature applications. This can bedone with hollow wing tips at the outer edge of the turbine. Thisspecial tip design increases turbine efficiency and provides a path forcooling air to get through the blades. Turbine wing cooling can also beprovided by compressed air from the compressor side fed across thehousing to the back side of the turbine wheel. In addition, a heat pipecan be used to cool the turbine wheel and blades.

In addition, a torsional softening device can be used on the power path.Crankshaft energy or rotational mechanical energy from a propulsiontrain can be brought through a flex shaft or an impulse softening device(either spring loaded or flexing) in such a way that torque impulsesfrom the engine or propulsion train are removed without loss of thatenergy, before entering the housing. By not impacting the transmissionwith high torque spikes on the traction drive, the peak torquerequirement is reduced. By eliminating these torque spikes, tractiondrives are more reliable, because the traction requirements are limitedby the maximum torque on the system. By minimizing these torque spikeson the traction drives, the size and surface contact areas of thetraction drives can be minimized. Minimal surface contact areas maximizeefficiency of the system, and can still achieve the torque required fortransmitting the continuous power.

Alternatively, and in accordance with another embodiment, a variablespeed traction drive design with fixed displacement hydraulic pumps inplace of the shaft, belt or gear drive may be utilized. This makes thesystem easier to package, which could be especially useful on very bigengines having multiple turbochargers.

In a further embodiment, illustrated in FIG. 13, a secondsuper-turbocharger is run off one transmission as a way to get a higherpressure ratio, and as a way to get cooler intake temperatures by usinga second intercooler. This is possible with a fixed speed ratio betweenthe two super-turbochargers. The first super-turbocharger 1302 has anair intake conduit 1308 and compresses air, which is supplied to theengine from compressed air conduit 1310. Exhaust air conduit 1314receives exhaust gas from the engine to run the turbine of the firstsuper-turbocharger 1302. The exhaust gas exits the exhaust exit conduit1312. The first super-turbocharger 1302 is coupled to the secondsuper-turbocharger 1304 with a transfer gear 1306.

FIG. 14A illustrates another embodiment of an implementation of the useof two super-turbochargers, such as low pressure super-turbocharger 1402and high pressure super-turbocharger 1404. A standard super-turbochargerdoes not do a good job of recovering the high-pressure pulse that comesout of the cylinder when the exhaust valve first opens. To improve thisimpulse pressure recovery, as illustrated in FIG. 14A, the high pressureexhaust valve ports 1406, 1408 are separated from the low pressureexhaust valve ports 1410, 1412 of a four-valve engine. The high pressureexhaust ports 1406, 1408 are directed to high pressure turbine 1434 viahigh pressure exhaust manifold 1430, while low pressure exhaust portsare directed to low pressure turbine 1420, via low pressure exhaustmanifold 1428. By changing valve timing of the valves in the highpressure exhaust ports 1406, 1408, such that valves on the high pressureexhaust ports 1406, 1408 are opened first and ported to the highpressure turbine 1434, the pulse energy is recovered better. The valveson the high pressure exhaust ports 1406, 1408 are closed quickly, andthen the valves on the low pressure exhaust ports 1410, 1412 are openedfor the duration of the exhaust stroke. The valves on the low pressureexhaust ports 1410, 1412 are ported to a low pressure turbine 1420. Thisprocess reduces the work required by the piston to exhaust the cylinder.This process improves idle fuel efficiency, or at least eliminatesparasitic losses at idle. The outlet of the high-pressure turbine 1434is also connected to the low-pressure turbine 1420. A catalyzed dieselparticulate filter (not shown) can also be disposed before the lowerpressure turbine.

As also illustrated in FIG. 14A, an EGR conduit 1438 is connected to thehigh pressure exhaust manifold 1430. The EGR conduit 1438 allows aportion of the exhaust from the high pressure exhaust manifold 1430 tobe channeled back to the intake manifold 1444, via cooler 1440 and EGRvalve 1442. The exhaust from the high pressure exhaust manifold 1430,that is channeled through the EGR conduit 1438, is channeled to theintake manifold 1444 for the purpose of the recirculation of exhaustgases. The exhaust gases flowing through the exhaust gas recirculatorconduit 1438 assist in lowering the combustion temperature in thecombustion chamber, especially after being cooled in cooler 1440. Theexhaust gases contain moisture and other liquids that assist in loweringthe temperature of the combustion chamber to thereby reduce NOxemissions from the engine. The amount of recirculated exhaust gas iscontrolled by the EGR valve 1442. EGR valve 1442 can be fixed, such asthrough the use of a restrictor valve, or can be varied, depending uponthe monitored NOx emissions of the engine.

As also shown in FIG. 14A, high pressure air is funneled through thehigh pressure compressor manifold 1446 from the high pressure compressor1432 to the intake manifold 1444. Hence, the intake manifold 1444 ismaintained at a predetermined high pressure level dictated by the outputof the high pressure compressor 1432. In order for the recirculatedgases to flow through the EGR conduit 1438, the pressure in the highpressure manifold 1430 must be higher than the pressure in the intakemanifold 1444, as dictated by the output pressure of the high pressurecompressor 1432. In that regard, the valves in the high pressure exhaustports 1406, 1408 are opened sufficiently early during the downstroke ofthe piston, when residual pressure still exists in the piston to createa sufficiently high pressure in the high pressure exhaust manifold 1430to drive the exhaust gases from the high pressure exhaust manifold 1430through the EGR conduit 1438. As disclosed below, the valves in the highpressure exhaust ports 1406, 1408 open at a point at which there is asmall amount of energy loss in the process of driving the pistonsdownwardly. The opening point of the high pressure valves is prior tobottom dead center, but beyond the point of maximum torque of the pistonon the crankshaft, which is the point at which the rods are atsubstantially 90°. This point occurs at approximately 100°. The amountof torque is proportional to the cosine of the angle of the rods, sothat the lower the piston is when the high pressure valves open, theless energy that is lost in driving the pistons. However, there is asubstantial amount of residual pressure left in the cylinder chamber,which can be exhausted from the cylinder chamber by the high pressurevalves prior to reaching bottom dead center, that can be used to drivethe exhaust gases in the EGR conduit 1438 into the high pressure turbine1434. By pre-exhausting the cylinder, using the high pressure valves ofthe high pressure exhaust ports 1406, 1408, a large amount of theresidual pressure in the cylinder is exhausted prior to opening of thelow pressure exhaust ports 1410, 1412. When opened, the low pressureexhaust ports 1410, 1412 are capable of exhausting most of the pressurefrom the cylinders. In this manner, the residual pressure in thecylinders is used to channel exhaust gas through both the EGR conduit1438, to reduce NOx emissions and to drive the high pressure turbine1434, which adds additional power and efficiency to the engine.

As also shown in FIG. 14A, the exhaust gases from the low pressureexhaust manifold are used to drive a low pressure turbine 1420 of thelow pressure superturbocharger 1402. Exhaust gases emitted by the highpressure turbine 1434 are combined with the low pressure exhaust gasesfrom the low pressure exhaust ports 1410, 1412 to drive the low pressureturbine 1420. Exhaust gases from the low pressure turbine 1420 areexhausted by exhaust outlet 1436. The low pressure turbine 1420 iscoupled to the low pressure compressor 1418, which compresses the inletair 1422 by a predetermined amount. Conduit 1424 channels the compressedair from the low pressure compressor 1418 to the input of the highpressure compressor 1432, which functions to further compress thepressurized air in 1424 to produce higher pressure compressed air, whichis channeled to the inlet manifold 1444 by high pressure compressormanifold 1446.

FIG. 14B illustrates a variation of the embodiment illustrated in FIG.14A. As illustrated in FIG. 14B, the high pressure exhaust ports 1406,1408 are combined into a high pressure exhaust manifold that is coupledto the high pressure turbine 1434. In other words, all of the highpressure exhaust from the high pressure exhaust manifold 1430 is appliedto the high pressure turbine 1434 to drive the high pressure turbine1434, which in turn drives the high pressure compressor 1432. The highpressure compressor 1432 receives compressed air in conduit 1424 fromthe low pressure compressor 1418 of low pressure super-turbocharger 1402that compresses the inlet air 1422. The output of high pressurecompressor 1432 is fed to the input manifold 1444 via high pressurecompressor manifold 1446. The low pressure compressor 1418 is driven bythe low pressure turbine 1420 that is driven by the low pressure exhaustgases, in the low pressure exhaust manifold 1428, that are emitted bythe low pressure exhaust ports 1410, 1412. Exhaust gases from the lowpressure turbine 1420 are exhausted through exhaust outlet 1436. Thehigh pressure gases from the high pressure exhaust manifold 1430, thatdrive the high pressure turbine 1434, are coupled to the exhaust gasrecirculation (EGR) conduit 1426 and transmitted back to the intakemanifold 1444. The high pressure gases from the high pressure exhaustmanifold 1430, that drive the high pressure turbine 1434, are notsubstantially reduced in pressure and have a sufficiently high pressureto insert the exhaust gases from the EGR conduit 1426 into the intakemanifold 1444. FIG. 14B provides the greatest reduction in NOx gases,since essentially all of the exhaust gases from the high pressureexhaust manifold 1430 are recirculated to the intake manifold 1444.

As also illustrated in FIG. 14B, a waste gate 1448 may be utilized tobypass high pressure exhaust gases from the high pressure exhaustmanifold 1430 to the EGR conduit 1426. The high pressure exhaust gases,at times, may be too hot and/or may provide exhaust gases at a pressurethat will overdrive the high pressure turbine 1434. In that instance,the waste gate 1448 can be opened to feed a portion of the high pressureexhaust gas from the high pressure exhaust manifold 1430 directly to theEGR conduit 1426. In addition, an EGR valve 1450 may be added, whichconnects the EGR conduit 1426 to the low pressure exhaust manifold 1428.If a sufficient amount of exhaust gases are being fed through the EGRconduit 1426, a portion of those gases may be directed from the EGRconduit 1426 to the low pressure exhaust manifold 1428 via EGR valve1450. The excess gases from the EGR conduit 1426 can then be used to runthe low pressure turbine 1420 to add additional power to the engine byincreasing the intake manifold pressure 1444. Use of the EGR valve 1450provides an additional manner in which recirculated gases can berecovered to add additional power to the engine and increase theefficiency of the operation of the engine.

FIG. 14C illustrates another modification of the embodiments of FIGS.14A and 14B. As shown in FIG. 14C, inlet air 1422 is compressed by lowpressure compressor 1418. The compressed air from the low pressurecompressor 1418 is fed by conduit 1424 to the intake manifold 1444. Asalso illustrated in FIG. 14C, the second high pressure turbine is notutilized and all of the recirculation gas is recirculated from the highpressure exhaust ports 1406, 1408 via EGR conduit 1426 to the intakemanifold 1444. Exhaust gases from the low pressure exhaust ports 1410,1412 are combined in conduit 1428 to operate low pressure turbine 1420.The exhaust gases are then exhausted at exhaust outlet 1436. Hence, allof the blow down gases from the high pressure exhaust ports 1406, 1408are fed back into the intake manifold 1444 to create a large reductionin NOx gases. Alternatively, an EGR valve 1450 can be used to channel aportion of the exhaust gases in the EGR conduit 1426 to the low pressureexhaust manifold 1428, which adds further power to the low pressureturbine 1420 and reduces the amount of recirculated gases in the EGRconduit 1426. The EGR valve 1450 can be adjusted to adjust the amount ofexhaust gases that are fed from the EGR conduit 1426 to the low pressureexhaust manifold 1428. This process may be beneficial if a sufficientamount of exhaust gases are recirculated in the EGR conduit 1426 toreduce NOx output of the engine.

FIG. 14D is a graph of the valve lift, cylinder pressure and flow rateversus the piston position after top dead center. As shown in FIG. 14D,the cylinder pressure 1450 steadily decreases after top dead center, allthe way through the stroke of the piston. The lift of the high pressurevalve 1456 creates the high pressure flow 1452. The lift of the highpressure valve 1456 occurs around 100° rotation and creates a large blowdown surge of the high pressure flow 1452 that is exhausted through thehigh pressure exhaust ports 1406, 1408 (FIGS. 14A, 14B and 14C). Thelift of the low pressure valve is illustrated at curve 1454. The lowpressure valve lift creates the low pressure flow 1458 in the lowpressure exhaust ports 1410, 1412. As a result, the cylinder pressure1450 is further reduced in the cylinder.

FIG. 14E is a PV graph of the cylinder pressure versus the volume in thecylinder, as the piston moves downwardly and then upwardly in thecylinder. Near zero represents the top dead center, while 1 representsthe bottom dead center of the rotation of the cylinder. Two curves areshown in FIG. 14E. Curve 1464 represents the curve of the cylinderpressure versus the volume for an engine that does not employ the Rileycycle. Curve 1462 is a curve that illustrates the cylinder pressureversus volume in the cylinder for a Riley cycle device, such asillustrated in FIGS. 14A-C. At point 1466, the high pressure valve isopened on the Riley cycle device, as illustrated in FIGS. 14A-C, and thepressure is reduced. The area 1468, between points 1466, 1470, isrepresentative of the energy lost by opening the high pressure valve.However, as indicated in FIG. 14E, at point 1472, the pressure in theRiley cycle device falls below the pressure in a non-Riley cycle deviceand remains below the pressure of the non-Riley cycle device all the waythrough to point 1474. Between 1472 and point 1474, there is lesspressure in the cylinder, which results in less backpressure on thecylinder as the cylinder moves from point 1472 to point 1474. The largeamount of area between the Riley cycle curve 1462 and the normal curve1464, between points 1472 and 1476, as indicated by 1478, is indicativeof the energy saved by movement of the piston in the cylinder at thelower pressure.

In an alternate embodiment, a super-turbocharger may be used as an airpump for after treatment, as well as for the engine and eliminates theneed for a separate pump just for the burner.

In another embodiment, a governor (not shown) is provided to preventover-speeding, keeping the compressor out of a surge condition andcontrolling to the maximum efficiency of the turbine and compressor. Asuper-turbocharger can be unique from a normal turbocharger because thepeak of the turbine efficiency and the peak of the compressor efficiencycan be at the same speed. Controlling to this peak efficiency speed fora given boost requirement can be modeled and programmed into anelectronic governor. An actuator can provide governing, although anactuator is not needed for the electric transmission.

In another embodiment, the oiling system for the super-turbochargerpulls a vacuum inside the housing, and therefore reduces aerodynamiclosses of the high speed components.

In another alternate embodiment, a dual clutch super-turbochargerincludes an automatically shifted manual transmission. This type oftransmission shifts very smoothly because it has a clutch on both ends.FIG. 3C illustrates that the transmission could be of many differenttypes.

In another embodiment, traction drives for both the transmission and thespeed reduction from the turbo shaft are used. With ball bearings, thetraction fluid works as the lubricant as well. During supercharging, thesystem improves load acceptance, reduces soot emissions, provides up to30% increase in low end torque and up to 10% increase in peak power.During turbo-compounding, the system provides improved fuel economy ofup to 10% and controls backpressure. For engine downsizing, the systemprovides 30% more low end torque that allows the engine to be 30 to 50%smaller, having lower engine mass and improved vehicle fuel economy of17% or more. FIG. 15 illustrates the simulated BSFC improvement for anatural gas engine.

Also, a catalyst, a DPF or even a burner plus DPF can be positioned infront of the turbine of the super-turbocharger to heat the exhaust gasto a higher temperature than the heat of the engine. Higher temperaturesexpand the air even further making the flow rate across the turbinehigher. Approximately 22% of this heat addition can be turned intomechanical work across the super-turbocharger, assuming 80% turbineefficiency. Normally, higher volume in the exhaust that is fed to theturbine would slow the turbine response and create even bigger turbolag, but the super-turbocharger overcomes this problem with the tractiondrive 114 and continuously variable transmission 116 driving thepressure response. Similar techniques using a catalytic converter aredisclosed in International Patent Application No. PCT/US 2009/051742filed 24 Jul. 2009 by Van Dyne et al. entitled “Improving FuelEfficiency for a Piston Engine Using a Super-Turbocharger” which isspecifically incorporated herein by reference for all that it disclosesand teaches.

FIG. 16 is a simplified single line form illustration of one embodimentof a high efficiency, super-turbocharged engine system 1600. As willbecome apparent to those skilled in the art from the followingdescription, such a super-turbocharged engine system 1600 findsparticular applicability in diesel engines and some spark ignited,gasoline engines that are used in passenger and commercial vehicles, andtherefore the illustrative examples discussed herein utilize such anenvironment to aid in the understanding of the invention. However,recognizing that embodiments of system 1600 have applicability to otheroperating environments such as, for example, land based, powergeneration engines, and other land based engines, such examples shouldbe taken by way of illustration and not by way of limitation.

As shown in FIG. 16, the super-turbocharger 1604 includes a turbine1606, a compressor 1608, and a transmission 1610 that is coupled to thecrank shaft 1612 of the engine 1602 or other portions of the propulsiontrain. While not required in all embodiments, the illustrated embodimentof FIG. 16 also includes an intercooler 1614 to increase the density ofthe air supplied to the engine 1602 from the compressor 108 to furtherincrease the power available from the engine 1602.

Super-turbochargers have certain advantages of turbochargers. Aturbocharger utilizes a turbine that is driven by the exhaust of theengine. This turbine is coupled to a compressor which compresses theintake air that is fed into the cylinders of the engine. The turbine ina turbocharger is driven by the exhaust from the engine. As such, theengine experiences a lag in boost when first accelerated until there isenough hot exhaust to spin up the turbine to power a compressor, whichis mechanically coupled to the turbine, to generate sufficient boost. Tominimize lag, smaller and/or lighter turbochargers are typicallyutilized. The lower inertia of the lightweight turbochargers allows themto spin up very quickly, thereby minimizing the lag in performance.

Unfortunately, such smaller and/or lighter weight turbochargers may beover-sped during high engine speed operation when a great deal ofexhaust flow and temperature is produced. To prevent such over speedoccurrences, typical turbochargers include a waste gate valve that isinstalled in the exhaust pipe upstream of the turbine. The waste gatevalve is a pressure operated valve that diverts some of the exhaust gasaround the turbine when the output pressure of the compressor exceeds apredetermined limit. This limit is set at a pressure that indicates thatthe turbocharger is about to be over-sped. Unfortunately, this resultsin a portion of the energy available from the exhaust gases of theengine being wasted.

Recognizing that conventional turbochargers sacrifice low endperformance for high end power, devices known as super-turbochargerswere developed. One such super-turbocharger is described in U.S. Pat.No. 7,490,594 entitled “Super-Turbocharger,” issued Feb. 17, 2009, whichis specifically incorporated herein by reference for all that itdiscloses and teaches.

As discussed in the above-referenced application, in asuper-turbocharger the compressor is driven by the engine crank shaftvia a transmission that is coupled to the engine during low engine speedoperation when sufficiently heated engine exhaust gas is not availableto drive the turbine. The mechanical energy supplied by the engine tothe compressor reduces the turbo lag problem suffered by conventionalturbochargers, and allows for a larger or more efficient turbine andcompressor to be used.

The super-turbocharger 1604, illustrated in FIG. 16, operates to supplycompressed air from the compressor 1608 to the engine 1602 withoutsuffering from the turbo-lag problem of a conventional turbocharger atthe low end and without wasting energy available from the engine exhaustgas heat supplied to the turbine 1606 at the high end. These advantagesare provided by inclusion of the super-turbocharger transmission 1610that can both extract power from, and supply power to, the engine crankshaft 1612 to both drive the compressor 1608 and load the turbine 1606,respectfully, during various modes of operation of the engine 1602.

During start up, when conventional turbochargers suffer a lag due to thelack of sufficient power from the engine exhaust heat to drive theturbine, the super-turbocharger 1604 provides a supercharging actionwhereby power is taken from the crank shaft 1612 via thesuper-turbocharger transmission 1610 to drive the compressor 1608 toprovide sufficient boost to the engine 1602. As the engine comes up tospeed and the amount of power available from the engine exhaust gas heatis sufficient to drive the turbine 1606, the amount of power taken fromthe crank shaft 1612 by the transmission 1610 is reduced. Thereafter,the turbine 1606 continues to supply power to the compressor 1608 tocompress the intake air for use by the engine 1602.

As the engine speed increases, the amount of power available from theengine exhaust gas heat increases to the point where the turbine 1606would over speed in a conventional turbocharger. However, with thesuper-turbocharger 1604, the excess energy provided by the engineexhaust gas heat to the turbine 1606 is channeled through thetransmission 1610 to the engine crank shaft 1612 while maintaining thecompressor 1608 at the proper speed to supply the ideal boost to theengine 1602. The greater the output power available from the exhaust gasheat of the engine 1602, the more power generated by the turbine 1606that is channeled through the transmission 1610 to the crank shaft 1612while maintaining the optimum boost available from the compressor 1608.This loading of the turbine 1606 by the transmission 1610 prevents theturbine 1606 from over speeding and maximizes the efficiency of thepower extracted from the engines exhaust gases. As such, a conventionalwaste gate is not required.

While the amount of power available to drive the turbine 1606 in aconventional super-turbocharged application is limited strictly to theamount of power available from the engine exhaust, the turbine 1606 iscapable of generating significantly more power if the thermal energy andmass flow supplied to the turbine blades can be fully utilized and/orcan be increased. However, the turbine 1606 cannot operate above acertain temperature without damage, and the mass flow is conventionallylimited to the exhaust gases coming out of the engine 1602.

Recognizing this, the embodiment of the system 1600 protects the turbine1606 from high temperature transients by placing a catalyzed dieselparticulate filter 1616 upstream of the turbine 1606. In one embodiment,the catalyzed diesel particulate filter is placed upstream from theturbine near the exhaust manifold which enables exothermic reactionsthat result in an increase in exhaust gas temperature during sustainedhigh speed or load operation of the engine. Using a catalyzed digitalparticulate filter, energy can be recovered from the soot, hydrocarbonsand carbon monoxide that is burned on the catalyzed diesel particulatefilter 1616 to add power to the super turbo charger which is locateddownstream from the catalyzed digital particulate filter 1616. Energyrecovery can be achieved from either a conventional diesel particulatefilter that has a very restricted flow-through capacity, with nearly100% soot collection, or by using a flow-through catalyzed digitalparticulate filter. A flow-through catalyzed digital particulate filteris a diesel particulate filter that only collects about half of the sootand lets the other half pass through. Both types of digital particulatefilters are catalyzed in order to have emissions burn at a reasonablylow temperature. Catalyzing of the digital particulate filter isaccomplished by providing a platinum coating to the particulate filterelements that ensures that soot, hydrocarbons and carbon monoxide burnat low temperatures. Additionally, it is possible to use a dieselparticulate filter and a burner to burn the soot off of the digitalparticulate filter upstream from the super-turbocharger. Gasolineengines typically do not have enough soot to require a dieselparticulate filter. However, some gasoline direct injection enginesproduce sufficient soot and other particulates so that the use of aparticulate filter may be beneficial, and the use of a catalyzed dieselparticulate filter may be deployed in the manner disclosed herein.

To cool the exhaust gas, prior to reaching the turbine, a portion of thecompressed air generated by the compressor is fed directly into theexhaust upstream from the turbine, via a control valve 1618, and addedto the engine exhaust gases leaving the catalyzed diesel particulatefilter 1616. The cooler intake air expands and cools the exhaust gas andadds additional mass to the exhaust gas flow, which adds additionalpower to the turbine 1606 as described in more detail below. As morecooler air is provided to the hot exhaust gases to maintain thetemperature of the combined flow to the turbine 1606 at the optimumtemperature, the energy and the mass flow that is delivered to theturbine blades also increases. This significantly increases the powersupplied by the turbine to drive the engine crank shaft.

So as to not interfere with the stoichiometric reaction within thecatalyzed diesel particular filter 1616, the compressor feedback air isadded downstream of the catalyzed diesel particulate filter 1616. Insuch an embodiment, the engine exhaust gas is passed through thecatalyzed diesel particulate filter 1616 and temperature of the exhaustgas is increased by the exothermic reaction. The compressed feedback airis then added and expands so that the total mass flow supplied to theturbine is increased. Embodiments of the present invention control theamount of compressed feedback air supplied to cool the exhaust and todrive the turbine to ensure that the combination of the coolercompressed feedback air and the engine exhaust gases are delivered tothe turbine at an optimum temperature for turbine blade operation.

Since the catalyzed diesel particulate filter 1616, illustrated in FIG.16, has a large thermal mass than the exhaust gases from engine 1602,the catalyzed diesel particulate filter 1616 operates as a thermaldamper initially, which prevents a high temperature thermal spike fromreaching the turbine 1606. However, since the reactions in the catalyzeddiesel particulate filter 1616 are exothermic in nature, the temperatureof the exhaust gases leaving the catalyzed diesel particulate filter1616 are higher than that of the exhaust gas entering the catalyzeddiesel particulate filter 1616. So long as the temperature of theexhaust gas entering the turbine remains below the maximum operatingtemperature of the turbine 1606, there is no problem.

However, during sustained high speed and high load operation of theengine 1602, the exit temperatures of the converted exhaust gas fromcatalyzed diesel particulate filter 1616 can exceed the maximumoperating temperature of turbine 1606. As set forth above, thetemperature of the exhaust gases exiting the catalyzed dieselparticulate filter 1616 are reduced by supplying a portion of thecompressed air from the compressor 1608 via a feedback valve 1618, andmixed with the exhaust gas exiting the catalyzed diesel particulatefilter 1616. Significantly improved fuel economy is achieved by notusing fuel as a coolant during such conditions, as is done inconventional systems. Additionally, the operation of the transmission iscontrolled to allow the compressor 1608 to supply a sufficient amount ofcompressed air to provide optimum boost to the engine 1602 and thecompressed feedback air to the turbine 1606 via the feedback valve 1618.The excess power generated by the turbine 1606 resulting from theincreased mass flow of the compressed air through the turbine ischanneled via the transmission 1610 to the crank shaft 1612, yet furtherincreasing fuel efficiency.

The output temperature of the compressed air from the compressor 1608 istypically between about 200° C. to 300° C. A conventional turbine canoperate optimally to extract power from gases at approximately 950° C.,but not higher without distortion or possible failure. Because of thematerial limits of the turbine blades, the optimal power is achieved atapproximately 950° C. Since the materials limit the exhaust gastemperatures to about 950° C., supplying more air to increase the massflow across the turbine at the temperature limit, e.g., 950° C.,increases the performance of the turbine.

While such a flow of compressed feedback air at 200° C. to 300° C. ishelpful in reducing the temperature of the exhaust gas coming out of thecatalyzed diesel particulate filter 1616, it is recognized that maximumpower from the turbine 1606 can be supplied when the temperature and themass flow is maximized within the thermal limits of the turbine 1606. Assuch, in one embodiment, the amount of feedback air is controlled sothat the combination of exhaust gas and feedback air is maintained at ornear the turbine's maximum operating temperature so that the amount ofpower delivered to the turbine is maximized or significantly increased.Since all of this excess power is normally not required by thecompressor 1608 to supply the optimum boost to engine 1602 and to supplythe compressor feedback air via feedback valve 1618, the excess powermay be transferred by the transmission 1610 to the crank shaft 1612 ofthe engine 1602 to thereby increase the overall efficiency or power ofthe engine 1602.

As discussed above, in one embodiment, the connection of the compressorfeedback air via feedback valve 1618 employs a catalyzed dieselparticulate filter 1616 as the thermal buffer between the engine 1602and turbine 1606. As such, the supply of air from the compressor isprovided downstream of the catalyzed diesel particulate filter 1616 soas to not disrupt the stoichiometric reaction within the catalyzeddiesel particulate filter 1616. That is, in embodiments that utilize acatalyzed diesel particulate filter 1616, supplying the compressorfeedback air upstream of the catalyzed diesel particulate filter 1616would result in excess oxygen being supplied to the catalyzed dieselparticulate filter 1616, thereby preventing the catalyzed dieselparticulate filter 1616 from generating a stoichiometric reaction thatis required for proper operation.

Since optimum efficiency of power generation by the turbine 1606 isachieved when the temperature of the gas mixture of the compressorfeedback air and exhaust gas on the turbine blades is maximized (withinthe material limits of the turbine itself), the amount of compressorfeedback air admitted by the feedback valve 1618 is limited so as to notreduce the temperature significantly below such an optimizedtemperature. As the catalyzed diesel particulate filter 1616 producesmore thermal energy via an exothermic reaction and the temperature ofthe converted exhaust gases from the catalyzed diesel particulate filter1616 increases to a temperature above the maximum operating temperatureof the turbine 1606, more compressor feedback air may be supplied viafeedback valve 1618 which increases the mass flow and energy supplied tothe turbine 1606. As the amount of thermal energy generated by catalyzeddiesel particulate filter 1616 is reduced, the amount of compressorfeedback air supplied by feedback valve 1618 can also be reduced so asto avoid supplying more air than necessary, which results in themaintenance of the temperature of the gas mixture at the optimumoperating condition.

In another embodiment, the system utilizes the feedback valve 1618 forfeeding back the cooler compressor air into the exhaust ahead of theturbine at low speed, high load operating conditions to avoid surgingthe compressor. Compressor surge occurs when the compressor pressuregets high but the mass flow allowed into the engine is low as a resultof the engine turning at a slow rpm and not requiring much intake airflow. Surging (or aerodynamic stalling) of the compressor resulting fromlow airflow across the compressor blades causes the efficiency of thecompressor to fall very rapidly. In the case of a normal turbocharger,enough surge can stop the turbine from spinning. In the case of asuper-turbocharger it is possible to use power from the engine crankshaft to push the compressor into surge. Opening the feedback valve 1618allows a portion of the compressed air to feedback around the engine.This feedback flow brings the compressor out of surge and allows higherboost pressure to reach the engine 1602, thereby allowing the engine1602 to generate more power than would normally be possible at lowengine speeds. Injecting the compressed air into the exhaust ahead ofthe turbine conserves the total mass flow through the compressor so thatall the flow reaches the turbine which minimizes the power needed fromthe engine to supercharge to a high boost pressure level.

In another embodiment, an additional cold start control valve 1620 maybe included for operation during rich engine cold starts. During such anengine cold start, the exhaust gases from the engine 1602 typicallyinclude excess un-burnt fuel. Since this rich mixture is notstoichiometric, the catalyzed diesel particulate filter 1616 is unableto fully reduce the un-burnt hydrocarbons (UHC) in the exhaust gas.During such times, the cold start control valve 1620 may be opened toprovide compressor feedback air to the input of the catalyzed dieselparticulate filter 1616 to supply the extra oxygen necessary to bringthe rich mixture down to stoichiometric levels. This allows thecatalyzed diesel particulate filter 1616 to light off faster and moreefficiently reduce the emissions during the cold start event. If theengine is idling, a normal turbocharger would have no boost pressure tobe able to supply the feedback air. However, the transmission ratio oftransmission 1610 can be adjusted to give enough speed to the compressorto generate the pressure needed for the air to flow through valve 1620.In that regard, control signal 1624 can be used to adjust the ratio oftransmission 1610 so that sufficient rotational speed can be providedfrom the engine drive shaft 1612 to the compressor 1608 during idling,especially during a cold start, to compress enough air to flow throughthe cold start valve 1620 and ignite catalyzed diesel particulate filter1616 with a sufficient amount of oxygen.

The requirement for the additional oxygen is typically limited in a coldstart event, and often lasts only for 30 to 40 seconds. Many vehiclescurrently include a separate air pump to supply this oxygen during thecold start event, at significant cost and weight compared to the limitedamount of time that such an air pump is required to operate. Byreplacing the separate air pump with the simple cold start control valve1620, significant costs, weight and complexity savings are realized.Because the super-turbocharger 1604 can control the speed of thecompressor 1608 via the transmission 1610, the cold start control valve1620 may comprise a simple on/off valve. The amount of air suppliedduring the cold start event can then be controlled by controlling thespeed of the compressor 1608 via transmission 1610 under operation ofthe control signal 1624.

The cold start control valve 1620 may also be used during periods ofextremely high temperature operation if fuel is used as a coolant withinthe engine and/or for the catalyzed diesel particulate filter 1616,despite the negative effect on fuel efficiency. In such situations, thecold start control valve 1620 will be able to supply the extra oxygennecessary to bring the rich exhaust back down to stoichiometric levelsto allow the catalyzed diesel particulate filter 1616 to properly reducethe unburned hydrocarbon emissions in the exhaust. This provides asignificant benefit to the environment over prior systems.

In embodiments where the cold start control valve 1620 is an on/offvalve, the system can modulate cold start control valve 1620 to vary theamount of compressed air supplied so as to bring the exhaust down tostoichiometric levels. Other types of variable flow control valves mayalso be used to accomplish this same function.

FIG. 16 also discloses a controller 1640. Controller 1640 controls theoperation of the feedback valve 1618 and the cold start valve 1620.Controller 1640 operates to optimize the amount of air flow throughfeedback valve 1618 for different conditions. The amount of air thatflows through the feedback valve 1618 is the minimal amount of air flowthat is necessary to obtain a specific desired condition, as describedabove. There are two specific conditions in which controller 1640operates feedback valve 1618, which are: 1) surge limit of thecompressor for a given boost requirement is proximate at low rpm, highload of the engine; and, 2) temperature of the gas mixture is proximateentering the turbine 1606 at high rpm, high load conditions.

As shown in FIG. 16, controller 1640 receives the gas mixturetemperature signal 1630 from a temperature sensor 1638 that detects thetemperature of the gas mixture of the cooling air supplied from thecompressor 1608 that is mixed with the hot exhaust gases produced by thecatalyzed diesel particulate filter 1616. In addition, the controller1640 detects the compressed air intake pressure signal 1632 that isgenerated by the pressure sensor 1636 that is disposed in the conduit ofcompressed air that is supplied from the compressor 1608. Further, anengine speed signal 1626 and an engine load signal 1628 that aresupplied from the engine 1602 or a throttle are fed to the controller1640.

With respect to control of the temperature of the gas mixture that issupplied to the turbine 1606 at high speed, high load conditions,controller 1640 limits the temperature of the gas mixture to atemperature that maximizes the operation of the turbine 1606, withoutbeing so high as to damage the mechanisms of the turbine 1606. In oneembodiment, a temperature of approximately 925° C. is an optimaltemperature for the gas mixture to operate the turbine 1606. Once thetemperature of the gas mixture that is fed into the turbine 1606 beginsto exceed 900° C., the feedback valve 1618 is opened, to allowcompressed air from the compressor 1608 to cool the hot exhaust gasesfrom the catalyzed diesel particulate filter 1616 prior to passing intothe turbine 1606. The controller 1640 can be designed to target atemperature of approximately 925° C., with an upper bound of 950° C. anda lower bound of 900° C. The limit of 950° C. is one at which damage tothe turbine 1606 may occur using conventional materials. Of course, thecontroller can be designed for other temperatures, depending upon theparticular types of components and materials used in the turbine 1606. Aconventional proportional integral derivative (PID) control logic devicecan be used in the controller 1640 to produce these controlled results.

The benefit of controlling the temperature of the gas mixture thatenters the turbine 1606 is that the use of fuel in the exhaust to limitthe turbine inlet temperatures of the gas mixture is eliminated. Usingthe flow of the cooler compressed air to cool the hot exhaust gases fromthe catalyzed diesel particulate filter 1616 requires a large amount ofair, which contains a large mass to achieve the desired coolertemperatures of the gas mixture. The amount of air that is required tocool the hot exhaust gases from the catalyzed diesel particulate filter1616 is large because the cooler compressed air from the compressor 1608is not a good coolant, especially when compared to liquid fuel that isinserted in the exhaust gas. The hot exhaust gases from the output ofthe catalyzed diesel particulate filter 1616 cause the cooler compressedgas from the compressor 1608 to expand to create the gas mixture. Sincea large mass of the cooler compressed air from the compressor 1608 isrequired to lower the temperature of the hot exhaust gases from thecatalyzed diesel particulate filter 1616, a large mass flow of the gasmixture flows across the turbine 1606, which greatly increases theoutput of the turbine 1606. The turbine power increases by thedifference of the power created by the differential of the mass flowminus the work required to compress the compressed air flowing throughthe feedback valve 1618. By obtaining the gas mixture temperature signal1630 from temperature sensor 1638 and controlling the addition ofcompressed air by feedback valve 1618, the maximum temperature is notexceeded.

Controller 1640 also controls the feedback valve 1618 to limit surge inthe compressor 1608. The surge limit is a boundary that varies as afunction of the boost pressure, the flow of air through the compressorand the design of the compressor 1608. Compressors, such as compressor1608, that are typically used in turbochargers, exceed a surge limitwhen the flow of intake air 1622 is low and the pressure ratio betweenthe intake air 1622 and the compressed air is high. In conventionalsuper-turbochargers, the flow of intake air 1677 is low when the enginespeed (rpm) 1626 is low. At low rpms, when the compressed air is notused in large volumes by the engine 1602, the mass flow of intake air1622 is low and surge occurs because the rotating compressor 1608 cannotpush air into a high pressure conduit without a reasonable flow ofintake air 1622. The feedback valve 1618 allows flow through thecompressed air conduit 1609 and prevents or reduces surge in thecompressor 1608. Once surge in the compressor 1608 occurs, the pressurein the compressed air conduit 1609 cannot be maintained. Hence, at lowrpm, high load operating conditions of the engine 1602, the pressure ofthe compressed air in the compressed air conduit 1609 may drop belowdesired levels. By opening the feedback valve 1618, the flow of intakeair 1622 through the compressor 1608 is increased, especially at lowrpm, high load operating conditions of the engine, which allows thedesired level of boost to be achieved in the compressed air conduit1609. Feedback valve 1618 can simply be opened until the desiredpressure in the compressed air conduit 1609 is reached. However, bysimply detecting boost pressure in the compressed air conduit 1609,surge will occur prior to the feedback valve 1618 being opened to bringthe compressor 1608 out of a surge condition.

It is preferable, however, to determine a surge limit and open thefeedback valve 1618 in advance, prior to the occurrence of a surgecondition. For a given rpm and desired boost level a surge limit can bedetermined. The feedback valve 1618 can begin to open prior to thecompressor 1608 reaching a calculated surge limit. Opening the valveearly allows the compressor to spool up to a higher boost pressure morequickly because the compressor stays closer to the higher efficiencypoints of the compressor operational parameters. Rapid boost pressurerise at low rpm can then be achieved. By opening the valve before surgeoccurs, a more stable control system can also be achieved.

Opening the feedback valve 1618 in such a way as to improve theresponsiveness of the engine 1602, is achieved by allowing the engine1602 to get to a higher boost pressure more quickly when the engine 1602is at a lower rpm. Compressor 1608 is also more efficient, which resultsin less work for the transmission 1610 to achieve supercharging. Surgelimit control can be modeled within standard model based controlsimulation code, such as MATLAB. Modeling in this manner will allowsimulation of the controller 1640 and auto-coding of algorithms forcontroller 1640.

A model based control system, such as described above, is unique, inthat the utilization of the transmission 1610 to control the rotation ofthe turbine 1606 and compressor 1608 generates boost pressure withoutturbo lag. In other words, the transmission 1610 can extract rotationalenergy from the crank shaft 1612 to drive the compressor 1608 to achievea desired boost in compressed air conduit 1609 very quickly and prior tothe turbine 1606 generating sufficient mechanical energy to drive thecompressor 108 at such a desired level. In this manner, controls in aconventional turbocharger to reduce lag are reduced or eliminated. Themodel based control of the controller 1640 should be designed tomaintain the optimum efficiency of the compressor 1608 within theoperational parameters of the compressor 108.

The control model of controller 1640 should also be carefully modeled onthe pressure operational parameters, as mapped against the mass flowallowed by the engine for a given target speed and load in which targetspeed and load may be defined relative to the position of the throttleof the vehicle. As shown in FIG. 16, the engine speed signal 1626 can beobtained from engine 1602 and is applied to the controller 1640.Similarly, the engine load signal 1628 can be obtained from the engine1602 and applied to controller 1640. Alternatively, these parameters canbe obtained from a sensors located on the engine throttle (not shown).The feedback valve 1618 can then be operated in response to a controlsignal 1642 generated by controller 1640. Pressure sensor 1636 generatesthe compressed air intake pressure signal 1632 that is applied to thecontroller 1640, which calculates the control signal 1642 in response toengine speed signal 1626, engine load signal 1628 and compressed airintake pressure signal 1632.

During operational conditions of the engine 1602, in which a surge limitis not being approached by the compressor 1608 and the temperature ofthe gas mixture, as detected by the temperature sensor 1638, is notreached, the feedback valve 1618 is closed so that the system works as aconventional super-turbocharged system. This occurs over a majority ofthe operating parameters of the engine 1602. When high load and low rpmconditions of the engine 1602 occur, the feedback valve 1618 is openedto prevent surge. Similarly, at high rpm, high load operating conditionsof engine 1602, high temperatures are produced in the exhaust gases atthe output of the catalyzed diesel particulate filter 1616, so that thefeedback valve 1618 must be opened to reduce the temperature of the fuelmixture applied to the turbine 1606 below a temperature which wouldcause damage to the turbine 1606.

FIG. 17 is a detailed diagram of the embodiment of the high efficiencysuper-turbocharged engine system 1600 illustrated in FIG. 16. As shownin FIG. 17, engine 1602 includes a super-turbocharger that has beenmodified, as described above with respect to FIG. 16, to provide overallhigher efficiency than conventional super-turbocharged engines, as wellas providing high, optimal efficiency in low rpm, high load operatingconditions, and high, optimal efficiency at high rpm, high loadconditions. The super-turbocharger includes a turbine 1606 that ismechanically connected by a shaft to compressor 1608. Compressor 1608compresses intake air 1622 and supplies the compressed intake air toconduit 1704. Conduit 1704 is connected to feedback valve 1618 andintercooler 1614. As disclosed above, intercooler 1614 functions to coolthe compressed air, which becomes heated during the compression process.The intercooler 1614 is connected to the compressed air conduit 1726which, in turn, is connected to the intake manifold (not shown) of theengine 1602. Pressure sensor 1636 is connected to the compressed airconduit 1704 to detect the pressure and supply a pressure reading viathe compressed intake air pressure signal 1632, which is applied tocontroller 1640. The feedback valve 1618 is controlled by a controllerfeedback valve control signal 1642 generated by the controller 1640, asdisclosed above. Under certain operating conditions, feedback valve 1618opens to supply compressed air from compressed air conduit 1704 to amixing chamber 1706.

As shown in the embodiment of FIG. 17, the mixing chamber 1706 simplycomprises a series of openings 1702 in the catalyzed diesel particulatefilter output conduit 1708, which is surrounded by the compressed airconduit 1704 so that compressed air supplied from the compressed airconduit 1704 passes through the openings 1702 to mix with the exhaustgases in the catalyzed diesel particulate filter output conduit 1708.Any desired type of mixing chamber can be used to mix the coolercompressed air with the exhaust gases to lower the temperature of theexhaust gases. Temperature sensor 1638 is located in the catalyzeddiesel particulate filter output conduit 1708 to measure the temperatureof the exhaust gases in the catalyzed diesel particulate filter outputconduit 1708. Temperature sensor 1638 supplies a gas mixture temperaturesignal 1630 to controller 1640, which controls the feedback valve 1618to ensure that the temperature of the exhaust gases in the catalyzeddiesel particulate filter output conduit 208 do not exceed a maximumtemperature that would damage to the turbine 1606. Catalyzed dieselparticulate filter 1616 is connected to the exhaust manifold 1710 by wayof catalyzed diesel particulate filter inlet conduit 1714. By locatingthe catalyzed diesel particulate filter 1616 proximate to the exhaustmanifold 1710, the hot exhaust gases from the engine flow directly intothe catalyzed diesel particulate filter 1616, which assists inactivating the catalyzed diesel particulate filter 1616. In other words,the proximate location of the catalyzed diesel particulate filter 1616near the outlet of the engine exhaust gases does not allow the exhaustgases to cool substantially prior to entering the catalyzed dieselparticulate filter 1616, which increases the performance of thecatalyzed diesel particulate filter 1616. As the exhaust gases passthrough the catalyzed diesel particulate filter 1616, the catalyzeddiesel particulate filter 1616 adds additional heat to the exhaustgases. These very hot exhaust gases at the output of the catalyzeddiesel particulate filter 1616 are supplied to the catalyzed dieselparticulate filter output conduit 208 and are cooled in the mixingchamber 1706 with the compressed intake air from the compressed airconduit 1704. Depending upon the temperature of the very hot exhaustgases that are produced at the output of the catalyzed dieselparticulate filter 1616, which varies depending upon the operatingconditions of the engine 1602, a different amount of compressed intakeair will be added to the exhaust gas during high speed, high loadconditions. During low engine speed, high engine load conditions, thefeedback valve 1618 also functions to allow intake air to flow throughthe compressor to avoid surge. Surge is similar to aerodynamic stall ofthe compressor blades, which occurs as a result of the low flowconditions through the compressor during low engine speed conditions.When surge occurs, the pressure in the intake manifold (not shown) fallsbecause the compressor 1608 is unable to compress the intake air. Byallowing air to flow through the compressor 1608 as a result of thefeedback valve 1618 being opened, pressure can be maintained in theintake manifold so that, when high torque is required at low enginespeeds, the high torque can be achieved because of the high intakemanifold pressure.

As disclosed above, when the engine is operating under high speed, highload conditions, the catalyzed diesel particulate filter 1616 causes alarge amount of heat to be generated in the exhaust gases that aresupplied to the catalyzed diesel particulate filter output conduit 1708.By supplying compressed, cooler intake air to the catalyzed dieselparticulate filter output conduit 1708, the hot exhaust gases under highspeed, high load conditions are cooled. As the load and speed of theengine increases, hotter gases are produced and more of the compressedair from conduit 1704 is required. If the turbine 1606 does not providesufficient rotational energy to drive the compressor, such as under lowspeed, high load conditions, the engine crank shaft 1612 can supplyrotational energy to the compressor 1608 via drive belt 1722, drivepulley 1718, shaft 1724, continuously variable transmission 1716 andtransmission 1728. Again, any portion of the propulsion train can beused to supply rotational energy to the compressor 1608, and FIG. 17discloses one implementation in accordance with one disclosedembodiment.

As also illustrated in FIG. 17, a cold start valve 1620 is alsoconnected to the compressed air conduit 1704, which in turn is connectedto the cold start conduit 1712. Cold start conduit 1712 is connected tothe catalyzed diesel particulate filter inlet conduit 1714, which isupstream from the catalyzed diesel particulate filter 1616. The purposeof the cold start valve is to provide compressed intake air to the inputof the catalyzed diesel particulate filter 1616 during startupconditions, as disclosed above. Under startup conditions, prior to thecatalyzed diesel particulate filter 1616 reaching full operationaltemperatures, additional oxygen is provided via the cold start conduit1712 to initiate the catalytic process. The additional oxygen that isprovided via the cold start conduit 1712 assists in the initiation ofthe catalytic process. Controller 1640 controls cold start valve 1620via controller cold start valve control signal 1644 in response to theengine speed signal 1626, engine load signal 1628, and the gas mixturetemperature signal 1630.

Hence, the high efficiency, super-turbocharged engine 1600 operates in amanner similar to a super-turbocharger, with the exception that feedbackvalve 1618 supplies a portion of the compressed air from the compressorto the input of the turbine for two reasons. One reason is to cool theexhaust gases prior to entering the turbine so that the full energy ofthe exhaust gases can be utilized and a waste gate is not needed underhigh speed, high load conditions. The other reason is to provide a flowof air through the compressor to prevent surge at low rpm, high loadconditions. In addition, the catalyzed diesel particulate filter can beconnected in the exhaust stream before the exhaust gases reach theturbine so that the heat generated by the catalyzed diesel particulatefilter 1616 can be used in driving the turbine 1606, and expanding thecompressed intake air that is mixed with the hot gases from thecatalyzed diesel particulate filter 1616, which greatly increasesefficiency of the system. Further, the cold start valve 1620 can be usedto initiate the catalytic process in the catalyzed diesel particulatefilter 1616 by providing oxygen to the exhaust gases during startupconditions.

Hence, a unique super-turbocharger is disclosed that uses a high speedtraction drive having a fixed ratio that reduces the rotationalmechanical speed of the turbine/compressor shaft to an rpm level thatcan be used by a continuously variable transmission that couples energybetween a propulsion train and the turbine/compressor shaft. Auniqueness of the super-turbocharger design is that the transmission isdisposed within the system. The continuously variable transmission isdisposed within a lower portion of the super-turbocharger housing. Thecontinuously variable transmission 1116 provides the infinitely variablespeed ratios that are needed to transfer rotational mechanical energybetween the super-turbocharger and the engine. Either a gearedcontinuously variable transmission can be used as continuously variabletransmission 1116 or a traction drive continuously variable transmissioncan be used. Hence, traction drives can be used for both the high speedtraction drive 114 and the continuously variable transmission 1116.

The foregoing description of the invention has been presented forpurposes of illustration and description. It is not intended to beexhaustive or to limit the invention to the precise form disclosed, andother modifications and variations may be possible in light of the aboveteachings. The embodiment was chosen and described in order to bestexplain the principles of the invention and its practical application tothereby enable others skilled in the art to best utilize the inventionin various embodiments and various modifications as are suited to theparticular use contemplated. It is intended that the appended claims beconstrued to include other alternative embodiments of the inventionexcept insofar as limited by the prior art.

What is claimed is:
 1. A super-turbocharger that is coupled to an enginesystem comprising: an engine; a turbine that generates turbinerotational mechanical energy from enthalpy of exhaust gas produced bysaid engine; a compressor that compresses intake air and suppliescompressed air to said engine; a shaft having portions that areconnected to said turbine and said compressor; a traction drive that iscoupled to said shaft that reduces rotational speed of said shaft to alower rotational speed at an output of said traction drive and transmitspower to and from said shaft; a transmission that is connected to saidoutput of said traction drive and said engine and that transmits saidpower between said engine system and said traction drive.
 2. Thesuper-turbocharger of claim 1 wherein said transmission comprises amechanical continuously variable transmission.
 3. The super-turbochargerof claim 1 wherein said transmission comprises a mechanical infinitelyvariable transmission.
 4. The super-turbocharger of claim 1 wherein saidtransmission comprises a discrete ratio mechanical transmission.
 5. Thesuper-turbocharger of claim 4 wherein said discrete ratio mechanicaltransmission comprises a fixed single ratio transmission.
 6. Thesuper-turbocharger of claim 4 wherein said discrete ratio mechanicaltransmission comprises a multiple-gear, shiftable transmission.
 7. Thesuper-turbocharger of claim 1 wherein said transmission contains aclutch to decouple said transmission from said engine system.
 8. Thesuper-turbocharger of claim 7 wherein said clutch comprises ahydrodynamic fluid coupling.
 9. The super-turbocharger of claim 1wherein said transmission comprises an electric motor-generator that iselectrically connected to said engine system.
 10. The super-turbochargerof claim 1 wherein said traction drive is a planetary traction drive.